The present invention is applicable to hydraulic actuators of sheet-bending presses, wherein there is need for displacing the rods of fluid-operated power cylinders that are interlinked to the crown of the press, and in dogless crown positioning. The invention is also applicable to hydraulic actuators for road-building and load-handling machines, as well for rolling mills.
Known in the prior art is an axial-piston pump with a valve directional control of the pressure fluid flow, said pump comprising a housing with an intake and pressure manifolds and a pressure fluid admission duct adapted to communicate with a source of pressure fluid. The pump pistons are oppositely arranged in axial borings in the housing with the capability of undergoing reciprocating motion and rotation around their axis. The pistons are adapted to interact with driving members secured to the drive shaft and forming pump delivery chambers. Serving as the driving members in the given particular pump construction are wobble plates set in position on the drive shaft for cooperatively rotating therewith.
Interworking of the pistons and wobble plates occurs due to the provision of a mechanical linkage for each of the pistons with said wobble plate, said mechanical linkage comprising a spherical head provided at the piston end, and a ball socket which is seated with its spherical surface on the spherical head. The end face of the ball sockets opposite that fitted with the piston heads, are forced against the wobble plate by means of a pressure disk loaded with springs mounted inside the housing. Each of the pump delivery chambers is connected to the pressure manifold through the delivery valve, and to the intake manifold through the inlet valve. The housing has radial borings accommodating plungers arranged coaxially with the delivery valve.
The plungers are adapted to establish, together with the boring in the pump housing, a chamber connected by a duct to the shaped surface of a cylindrical sleeve mounted in an axial boring of the pump housing for axial displacement, said sleeve embracing the drive shaft and being connected thereto for cooperatively rotating therewith.
The actuator spindle of the inlet valve is adapted to interact with the plunger. The outside surface of the cylindrical sleeve has depressions bounded by lands so as to form individual zones. One of said zones is constantly in communication with the exhaust line, while the other is in communication with the pressure fluid admission duct adapted to communicate with the source of pressure fluid which establishes a pressure of about 15 kgf/cm.sup.2 effective within said zone. The pressure of the fluid is transmitted through a duct for application to the plunger, and the latter causes the inlet valve to open and maintains it in that position. The period of time during which the inlet valve is kept open is defined by the axial position of the shaped sleeve which is acutated by a tie-rod with one end pin-coupled to the sleeve, while its free end projects outside the pump housing so as to be associated with any of the heretofore known mechanisms capable of displacing the tie-rod and fixing it in a preset position.
When the pump operates at the rated delivery, a positive opening of the inlet valve occurs at the instant when the pistons are commencing the admission stroke, that is, they start moving from the top dead center towards the bottom dead center, whereas closing of the inlet valve occurs at the moment when the pistons start the return stroke. When adjusting the pump delivery at the expense of axial displacement of the shaped sleeve, the inlet valve is kept positively open during a portion of the stroke of the pistons from the bottom dead center towards the top dead center, i.e., over a part of the discharge stroke. As a result, the piston while running, expels the pressure fluid from the delivery chamber into the intake manifold.
Rotation of the drive shaft is imparted to the wobble plates to cause the pistons to reciprocate. When the pistons perform the admission stroke, that is, the volume of the delivery chambers increases, pressure fluid is admitted to pass from the intake manifold through the open inlet valve to the delivery chambers so as to fill the displacement volume vacated by the pistons. Just before the commencing of the admission stroke, the inlet valve is positively opened by the plunger by virtue of the fluid pressure exerted thereupon, and the inlet valve remains open throughout the admission stroke. This is conducive to lower hydraulic pressure losses of the valve during admission, whereby the suction capacity and operating rate of the pump are increased. Upon completion of the admission stroke, the inlet valve is closed by the action of the spring with which it is loaded, in response to relieving the fluid pressure. The pistons perform the return motion, i.e., decreasing the volume of the delivery chambers (discharge stroke), and the entire pressure fluid displaced by the pistons, flows through the delivery valve to the pressure manifold. Thus, the pump operates at maximum delivery. The fluid pressure is appplied to the plunger through the shaped surface of the cylindrical sleeve, viz., through the zone in communication with the fluid. The fluid pressure applied to the plunger is relieved likewise through the shaped surface of the cylindrical sleeve, namely, through the zone constantly in communication with the exhaust line.
The inlet valve can be kept open within any portion of the discharge stroke depending upon the preset axial position of the shaped cylindrical sleeve with respect to the pump housing. Thus, the pump produces one-half its rated delivery when the inlet valve is closed mid-way of the discharge stroke, i.e., the pressure fluid displaced by the pistons during the first half of the discharge stroke is fed to the intake manifold, and during the second half of the discharge stroke, to the pressure manifold.
When the inlet valve is not closed altogether throughout the discharge stroke, the entire pressure fluid is discharged by the piston to the intake manifold, and the pump produces zero delivery, accordingly.
The operating rate of the pump, i.e., the period of time within which the maximum delivery of the pump can be reduced to zero, is dependent upon the travelling speed of the shaped cylindrical sleeve. The low weight of the sleeve, its free travel and its relatively short stroke (20 to 25 mm) make it possible to effect sleeve displacement within a reasonably short span of time without any substantial energy expense or major constructive difficulties. The operating rate of the pump is directly proportional to the speed of travel of the control sleeve up to a certain point, whereupon further increase in the sleeve displacement speed will not result in a higher pump operating rate.
The maximum operating rate depends upon the drive shaft rotational speed and is sufficiently high over an effective range of shaft speeds.
Thus, for instance, the operating rate of the pump equals 0.03 s at 1000 rpm of the drive shaft, and 0.02 s at 1500 rpm.
When the shaped sleeve travels in a reverse direction, i.e., so as to increase the pump delivery from zero to maximum, the operating rate of the pump is practically unrestricted, being proportional to the travelling speed of the sleeve.
The high pump operating rate and its ready attainment make the afore-discussed prior-art pump applicable to great advantage in automatic control systems, high-speed hydraulic machines etc.
The above-described pump, however, suffers from the disadvantage that for displacing the shaped cylindrical sleeve from the position corresponding to zero delivery within a space of time shorter than 0.02 s, a considerable proportion of the pressure fluid is required, for example, for displacing the piston of the mechanism for setting the axial position of the cylindrical sleeve with respect to the pump housing.
Moreover, when the delivery chambers of the pistons are to be united into two or more groups so as to establish a split-flow pump, the shaped cylindrical sleeve which is in fact a pump delivery control device, provides during axial travel with respect to the pump housing, only a concurrent alteration of the pump delivery in each of the pump delivery lines, since the sleeve shaped surface, viz., the zone communicating with the pressure fluid admission duct, acts through the respective ducts simultaneously upon all the plungers which operate the inlet valves.
Known in the art is a pump available from the firm "Sack & Kiesselbach", wherein the delivery chambers of the pistons are banked into two or more groups, each being in communication with the pressure manifold. The number of pressure manifolds corresponds to the number of the groups.
The above pump provides for both a concurrent variation of the delivery in all the pressure manifolds and an independent change of the delivery in each of the pressure manifolds.
A simultaneous change of the delivery in all the pressure manifolds is attained by a device which serves for a concurrent control over the delivery in each of the delivery chambers. Said device is essentially a variable orifice positioned in the pressure fluid admission line to the pump intake manifold.
The delivery chambers of the pistons are in communication with the pressure manifold through the delivery valves, and with the pressure manifold through the inlet valves. To effect an independent control of the amount of pressure fluid discharged by each of the pistons into the pressure manifold, each of the inlet valves is provided with a linkage actuator for a positive opening of the valve and for keeping it open over the discharge stroke. This serves to cut off one or a group of the pistons from operation for a preset period of time within which the delivery in one of the pressure manifolds drops to a required value.
Such a positive opening of the inlet valves substantially complicates the pump construction and, in addition, is applicable only to low-delivery pumps, e.g., under 20 to 25 l/min, since the control of delivery in a pump having a delivery in excess of 20 to 25 l/min by virtue of throttling the flow of pressure fluid admitted to the piston chambers, causes cavitation phenomena in the pistons, followed by destruction of the elements of the piston delivery chambers.
Hydraulic drives used to effect in-step traversing of two or more fluid-operated power cylinders (hydraulic motors) can be classified largely into three groups viz., synchronized hydraulic drives comprising devices capable of synchronizing the travelling speed of the cylinder rods; cophased hydraulic drives incorporating devices adapted to establish synchronism in the mutual position of the cylinder rods; and synchro-cophased hydraulic drives provided with devices ensuring synchronization of the travelling speed of the cylinder rods, and devices to effect synchronism in the mutual position of the cylinder rods.
In synchronized hydraulic drives with synchronization of the travelling speeds of the cylinder rods, the required accuracy is attained by metering the amount of fluid admitted to the fluid-operated power cylinders or hydraulic motors. In the synchronized hydraulic drives used currently, the metering is carried out by means of:
throttles or flow velocity governors mounted in the delivery or exhaust lines of fluid-operated power cylinders;
series-communicated pressure chambers of power cylinders;
special metering cylinders;
flow dividers (adders);
auxiliary control cylinders (hydraulic motors);
reference and dual pumps;
metering devices of hydraulic motors, or pumps operating as hydraulic motors.
Synchronized hydraulic drives are only capable of rough synchronism, since they fail to take account of the difference in the geometric volume of the fluid-operated power cylinders, the condition of their sealing elements, and the elasticity and temperature expansion of pressure fluid, the pipings and the power cylinders.
In cophased hydraulic drives, a continuous synchronization of the mutual position of the cylinder rods is carried out by means of:
a rigid mechanical linkage between fluid-operated power cylinders;
a feedback coupling to compensate for an error in the position of the cylinder rods with the use of servo-systems. Cophased hydraulic drives are mostly capable of a required accuracy of synchronism; however, hydraulic drives with a rigid mechanical coupling are bulky, whereas those with compensation of the cylinder rod position error by means of servo-systems suffer from considerable energy losses and a restricted range of operating speeds within which the hydraulic driver performs a stable operation free from self-vibrations.
Synchro-cophased hydraulic drives adapted to synchronize the travelling speed of the cylinder rods and their mutual position, are in fact a combination of several design versions of the synchronized and cophased hydraulic drives. The most commonly encountered version of a synchro-cophased hydraulic drive utilized in hydraulic presses is a combination of dual pumps and a feedback coupling to compensate for an error of the mutual position of the cylinder rods by virtue of an appropriate redistribution of pressure fluid fed by the pumps.
Known in the art is a synchro-cophased hydraulic drive for a sheet-bending press.
Said known hydraulic drive comprises fluid-operated power cylinders whose rods are interconnected with the crown of the press, a device for measuring the position error of the cylinder rods, a dual variable-delivery pump, a means to protect the delivery lines against overloads, and a source of pressure fluid.
The device for measuring the position error of the cylinder rods is made as a flexible linkage such as chain or wire rope, one of the ends of which is connected to the stationary press member, and the other end is held in position on one of the lever arms, while the other arm thereof is adapted to interact with the spool of a three-position triple-port servo-valve. The middle segment of the flexible linkage is adapted to interact with two sprockets or rollers that are free to rotate around a pivot secured to the press crown. The dual variable-delivery pump incorporates two pumps, each being connected with its drive shaft to the motor shaft through a common gear reducer.
Said dual pump has a device for varying its delivery which makes it possible for each pump to deliver into the pressure manifold practically the same amount of pressure liquid irrespective of the position assumed by the device.
The delivery line of each pump is communication, through the valve capable of keeping the crown in position upon ceasing the delivery of pressure fluid by the pump, with the inlet of a four-port directional control valve in communication with the source of pressure fluid. The latter in turn is in communication with a safety valve which is to maintain constant pressure in the delivery line of the source of pressure fluid.
One outlet line of the first four-port directional control valve in communication with the rod space of the first fluid-operated power cylinder, whereas the other outlet line is in communication with the head space of the second fluid-operated power cylinder. One outlet line of the second four-port directional control valve is in communication with the rod space of the second fluid-operated power cylinder, whereas the other outlet line is in communication with the head space of the first fluid-operated power cylinder. The fourth line of each directional control valve is in communication with the exhaust line. The means for protecting the delivery lines against overloading by the dual pump is fashioned as a servo-actuated safety valve which, when in the initial position, is normally open, and the delivery line of each pump is in communication with the exhaust line through the check valve and the normally open safety valve. When the fluid-operated power cylinder is operative, the safety valve is closed so as to maintain a preset operating pressure in the delivery lines of the dual pump.
One of the lines of the triple-port servo-valve is in communication with the exhaust line, while the other two lines are in communication with the delivery lines of the dual pump.
When the hydraulic drive operates, e.g., for the crown to travel from the top dead center towards the bottom dead center, the safety valve is closed, while the four-port directional control valves are so set that pressure fluid from the delivery lines of the dual pump is fed to the piston-head spaces of the fluid-operated power cylinders, while the piston-rod spaces of the latter are in communication through said directional control valves with the exhaust line. If the press crown runs askew due an erroneous positioning of the rods of the fluid-operated power cylinders, the flexible linkage deflects the lever arm with which it is interconnected, whereas the other arm of the lever causes the spool of the triple-port servo-valve to travel so as to bypass part of the pressure fluid passing from the dual pump delivery line, into the piston-head space of the advancing power cylinder and further to the exhaust line, thereby bringing the cylinder rods in synchronism.
In case of overloads, i.e., when the pressure in the dual pump delivery lines rises above a specified level, the servo-actuated safety valve lets the pump-delivered pressure fluid bypass to the exhaust line, thereby maintaining the pressure in the delivery lines within the preset limits.
The known hydraulic drive, however, sustains additional hydraulic losses stemming from throttling of the pressure fluid when passing through the triple-port servo-valve.
Moreover, said hydraulic drive is capable of but a restricted range of operating speeds of the cylinder rods, within which a stable traversing of the rods is ensured, with a broad range of effective pressures in the fluid-operated power cylinders. This is due to the fact that the servo-valve flow-rate characteristic curve is very steep at pressures approximating the upper limit of the effective pressure range. Due to such a steep curve, great acceleration of the cylinder rods occurs which results in overshooting of the synchronous position of the cylinder rods. Such over-shooting brings the cylinder rods into self-vibration which leads to imposition of heavy loads upon the press and eventually to impaired quality of sheet stock bending.
Furthermore, the known hydraulic drive fails to provide dogless stopping of the press crown when the latter approximates a preset position which involves the use of stop dogs adapted to take up the press force, whereby said stop dogs are made massive and bulky, thus adding to the weight of the press and taking much time to be readjusted for stopping the crown in another position.